The present invention pertains to the field of heating, ventilating, and air conditioning (xe2x80x9cHVACxe2x80x9d). More particularly, this invention relates to systems and methods for controlling the temperature and humidity of an enclosed space.
The quality of indoor air has been linked to many illnesses and has been shown to have a direct impact on worker productivity. New research strongly suggests that indoor humidity levels may have a significant impact on the health of building occupants. For example, microbes such as mold and fungus, which proliferate at higher indoor humidity levels, have been shown to emit harmful organic compounds. In addition to direct health effects, often the primary air quality complaint of building occupants is unpleasant odors associated with microbial activity. Building operators often attempt to eliminate odors by increasing outdoor air quantities. This usually exacerbates the problem because increasing outdoor air quantities often results in higher indoor air humidity levels, which, in turn, fosters further microbial activity.
The HVAC industry has responded to these indoor air quality (xe2x80x9cIAQxe2x80x9d) concerns through its trade organization, the American Society of Heating, Refrigerating and Air-Conditioning Engineers (xe2x80x9cASHRAExe2x80x9d). ASHRAE Standard 62-1999, Ventilation for Acceptable Indoor Air Quality, sets minimum ventilation rates and other requirements for commercial and institutional buildings. Meeting these standards generally requires systems capable of providing an increased supply of outdoor air to the conditioned space while maintaining acceptable humidity levels within the space. A large body of research supports the need for continuous ventilation in accordance with ASHRAE 62-1999, while maintaining the relative space humidity between 30% and 60%. IAQ problems including unacceptable odors and microbial infestation often occur when HVAC systems fail to meet these design criteria.
Commercial and institutional facilities often use xe2x80x9cpackagedxe2x80x9d units, which combine air conditioning, heating and sometimes air handling equipment in a single housing. Such systems are generally designed to provide inexpensive heating and cooling. Such packaged units are generally installed outside the building envelope, frequently at ground level or on the building roof. A typical packaged unit includes a supply fan and filter, a return air fan, a heating source (typically an indirect gas fired heater or electric heating coil), an outdoor air intake, and a mechanical refrigeration system consisting of a compressor, cooling coil, and a condensing coil with a fan that rejects heat to the outdoors. Typically a small fraction of outdoor air is mixed with a much larger fraction of return air from the building, conditioned by the unit then circulated through the building by means of a system of supply and return ductwork. The advantages of such packaged equipment include low purchase cost, simplicity, familiarity, and compact design. More than 80% of all air-conditioning systems sold to the commercial marketplace involve compressorized package equipment.
A significant shortcoming of such packaged HVAC units is that they are typically designed to utilize minimal outdoor air, and, as such, are frequently incapable of handling the increased continuous supply of outdoor air necessary to comply with ASHRAE 62-1999 guidelines. This is especially true in applications where the need for 100% outdoor air systems exist, such as makeup air to restaurants and hotel facilities. It is also true for applications like schools, movie theatres and other facilities where a high occupancy density results in the need for very high outdoor air percentages being provided by the HVAC system.
To meet the increased outdoor air requirements of the ASHRAE standards, HVAC professionals have attempted to use oversized packaged equipment to match the increased cooling load associated with higher outdoor air percentages. However, such oversized systems generally suffer from sub-par performance and are expensive to operate. As importantly, the oversized cooling capacity required to meet peak outdoor air load conditions proves excessive at the more common part-load conditions, and creates serious performance problems ranging from over-cooling the space and lost humidity control due to reduced compressor cycle times to freezing up coils and shortened compressor life. Therefore, providing outdoor air continuously presents a tremendous challenge to conventional packaged HVAC equipment.
For example, on mild, humid days (part-load conditions) an oversized packaged unit will quickly cool the space to a set temperature and then shut off the compressor. If the evaporator fan is kept running to maintain a continuous flow of outdoor air to the space, the indoor humidity level will usually climb due to the humidity level of the outdoor air being introduced. This increase in humidity will continue until the space temperature rises to the point that the thermostat once again calls for cooling. By this time, the humidity of the return air entering the cooling coil of the packaged HVAC system is elevated. The elevated humidity of the return air results in an elevated dew point temperature leaving the cooling coil. Typically, the system can maintain space temperature, but humidity control is lost, resulting in uncomfortable, cold, clammy conditions. Occupants will often respond by lowering the thermostat setting, causing the space relative humidity to further increase. If such high humidity conditions persist, microbial growth and other moisture-related IAQ problems may arise.
Another problem associated with oversized packaged equipment selected to process outdoor air on a continuous basis results from the re-evaporation of moisture that has condensed on the evaporator coil. Henderson et al. (1998) and Khattar et al (1985) both have confirmed the phenomenon, often observed in the field, where the actual moisture removed by a packaged HVAC unit is significantly less than anticipated based upon published performance data. Their research shows that this reduction in dehumidification capacity is attributable to moisture condensed on the direct expansion (DX) coil evaporating back into the supply air stream when the coil is cycled off but the fan continues to operate. Henderson (1998) has shown that evaporation of moisture condensed on the DX coil can reduce actual latent heat removal to less than 50% of the unit""s capacity at part load conditions. (1) Henderson, H. 1998. The Impact of Part Load Air Conditioner Operation on Dehumidification Performance: Validating a Latent Capacity Degradation Model. Proceedings ASHRAE IAQ 98. (2) Khattar, M et. al. 1985. Fan Cycling Effects on Air Conditioner Moisture Removal Performance in Warm, Humid Climates. Presented at the International Symposium on Moisture and Humidity, Proceedings. April, 1985, Washington D.C. (3) Henderson, H. 1990. An Experimental Investigation of the Effects of Wet and Dry Coil Conditions on Cyclic Performance in the SEER Procedure. Proceedings of USNC/IIR Refrigeration Conference at Purdue University, West Lafayette, Ind. July, 1990.) These and other limitations present significant problems when packaged rooftop systems are forced to handle high percentages of outdoor air volume, particularly if operated as 100% outdoor systems. When applying a conventional packaged rooftop system to handle all outside air, the cooling tons required at peak conditions are far greater than the cooling output available at the rated airflow of the conventional unit. This occurs because standard conventional packaged cooling equipment currently available on the marketplace by the major HVAC equipment manufacturers is generally designed to accommodate only a relatively small portion of outdoor air, typically 10-20%.
For example, a typical packaged gas/electric rooftop unit available on the market today may have a rated cooling performance at 95xc2x0 Fahrenheit (F) ambient, 80xc2x0 F. coil entering dry bulb, 67xc2x0 F. coil entering wet bulb in accordance with the ARI Standard 210/240-94. Assuming a typical ASHRAE/ARI outdoor air cooling design condition of 95xc2x0 F. dry bulb and a 78xc2x0 F. web bulb, and a return air condition of 78xc2x0 F. dry bulb and 50% relative humidity, the mixed air condition entering the cooling coil of 80xc2x0 F. and 67xc2x0 F. wet bulb corresponds to an approximately 12% outdoor air percentage based on a simple mixed air calculation.
Therefore, the design standard used to rate standard packaged cooling equipment assumes that 80-90% of the air delivered to the cooling coil is conditioned return air from the space. This return air stream requires far less cooling capacity to condition than raw outdoor air during peak cooling design conditions. As such, the total cooling capacity needed by the standard conventional packaged equipment would be greater if it were designed to accommodate a much higher percentage of outdoor air.
For example, conditioning a 1,500 cubic feet per minute (cfm) outdoor air stream from 85xc2x0 F. and 130 grains (enthalpy of 40.8 BTU/pound) to a 56xc2x0 F. dew point (enthalpy of 23.8 BTU/pound) requires approximately 10 tons of cooling capacity based on a simple psychrometric calculation ((1500 cfmxc3x974.5xc3x97(40.8-23.8)/12000 BTU/ton of cooling). However, the recommended minimum amount of air capacity that can be processed by a typical 10 ton unit (alternative 1) without potentially causing problems such as frosting and compressor failure is approximately 3,000 cfm (300 cfm/ton). If the unit is set up to provide 50% outdoor air (1,500 cfm), and 50% return air (1,500 cfm) for a total of 3,000 cfm across the cooling coil, the cooling capacity must be increased to a 15 ton (alternative 2) unit to accommodate the load associated with the extra 1,500 cfm of recirculated air. Problems such as coil frosting may be avoided in many cases, since the mixed air temperature to the cooling coil is much closer to the aforementioned design conditions of 80xc2x0 F. and 67xc2x0 F. wet bulb. Examples of alternatives 1 and 2 are presented below.
If a standard 10 ton system is operated with only 1500 cfm of air passing across the coil (only 150 cfm/ton), and if this air is all outdoor air, the full 10 tons of cooling will be required to reach a supply condition with a 56xc2x0 F. dew point. However, when the outdoor air drops from the peak design condition of 95xc2x0 F. and 78xc2x0 F. wet bulb to say 78xc2x0 F. and 64xc2x0 F. wet bulb, the 10 ton compressor will deliver air as cool as 30xc2x0 to 34xc2x0 F. At this point, the refrigeration pressure and temperature will be very low, low enough to cause the moisture condensed on the cooling coil to freeze. This frost buildup can result in increased pressure loss across the cooling coil, which results in a reduction of airflow, which results in more significant frost formation. This and other problems associated with operating conventional DX cooling systems at reduced airflow are well known to the industry and those skilled in the art of refrigeration.
By applying a 15 ton system to process a total of 3,000 cfm, 1,500 cfm of which is outdoor air with the remainder being return air, the mixed air condition to the coil is decreased from the 95xc2x0 F. and 78xc2x0 F. wet bulb mentioned in the previous example to approximately 86.5xc2x0 F. and 72xc2x0 F. web bulb. At the peak condition, the 15 tons will provide a supply condition having a dew point of approximately 56xc2x0 F. At the part load condition used previously, 78xc2x0 F. and 64xc2x0 F. wet bulb, the supply air condition will be approximately 40xc2x0 F. At this condition, the refrigerant temperature is not as cold as the previous example, and therefore may allow the coil to be operated without freezing under part load condition.
However, using the increased cooling tons and supply airflow may cause other operational problems. The higher 3,000 cfm supply airflow quantity may, for example, overcool the space, especially at part-load conditions. This cooling causes the compressor to cycle off, resulting in the delivery of high humidity air directly to the space in addition to the moisture evaporated from the cooling coil if the supply air fan continues to run. If, as a third alternative, a 10 ton unit is used to process the 3,000 cfm of total airflow, of which 1,500 cfm is outdoor air, at typical cooling season latent design condition of 85xc2x0 F. and 130 grains, most conventional packaged units of this size are only capable of delivering air at a dew point of approximately 59xc2x0 F., even at a favorable, return air condition of 75xc2x0 F. and 60% relative humidity, and would therefore be incapable of maintaining the space at the desired level of 50% relative humidity, since a dew point of approximately 55xc2x0 F. is required even if there is no latent load generated by people or infiltration.
Customized overcooling reheat systems have been used in an attempt to overcome these problems. However, such systems are expensive to purchase and operate. Furthermore, complicated refrigeration circuits frequently employed by such systems can be difficult to troubleshoot and expensive to maintain. An example of the complexity required to deliver a packaged piece of equipment to effectively condition outdoor air is the TRANE(copyright) FAU product recently introduced to the marketplace. The TRANE(copyright) Applications Considerations Bulletin MUA-PRC004-EN shows a system that includes two separate evaporator coils (an outdoor air evaporator and a main evaporator), three separate condensing coils (a reheat condenser, a reheat outdoor condenser and a main condenser), one reheat compressor, three main compressors, two expansion valves, a subcooler and multiple complex controls.
Another attempt to meet the outdoor air and humidity level requirements of the ASHRAE standards is through the use of xe2x80x9cactivexe2x80x9d desiccant-based systems, desiccant systems that employ a heated regeneration air stream to remove moisture from the air. These active desiccant systems have been used to reduce the humidity of outdoor air prior to its introduction to the conventional HVAC system or directly to the conditioned space. This allows the packaged equipment to better control the space humidity despite increased outdoor air requirements. Desiccants can be solid or liquid substances that have the ability to attract and hold relatively large quantities of water. In many commercial air conditioning applications where desiccants are used, the desiccant is in a solid form and absorbs moisture from the air to be conditioned. Examples of these types of desiccants are silica gel, activated alumina, molecular sieves, and deliquescent hygroscopic salts. In some cases, these desiccants are contained in beds over which the air to be conditioned is passed. Many times, however, the desiccant is contained in what is known as an xe2x80x9cactive desiccant wheel.xe2x80x9d
An active desiccant wheel is an apparatus typically comprising closely spaced, very thin sheets of paper, polymer film or metal which are coated or impregnated with a desiccant material. The wheel is usually contained in duct work or in an air handling system that is divided into two sections: a supply section and a regeneration section. The wheel is rotated slowly on its axis such that a given zone of the wheel is sequentially exposed to the two sections. In the supply section, the desiccant is contacted by the supply/outdoor air. In this section, the desiccant wheel dehumidifies the supply/outdoor air stream by absorbing moisture from the air onto its desiccant surface. In the regeneration section, the desiccant contacts a regeneration air stream (e.g., return/exhaust air being discharged from the space or raw outdoor air). This regeneration air desorbs the moisture from the desiccant that was adsorbed from the supply/outdoor air. A heater is often used to heat the regeneration air stream as needed to regenerate (i.e., dry) the desiccant wheel as it rotates through the regeneration air stream. By cycling the wheel through these two air streams, the adsorbing/desorbing operation of the wheel is continuous and occurs simultaneously.
In the past, most active desiccant preconditioning systems have not been coupled with rooftop packaged equipment, but applied as stand alone systems. When they have been coupled with rooftop packaged equipment, they have been positioned upstream of the packaged unit in an attempt to control the humidity of the air entering the conventional vapor compression system. Such systems have processed the outdoor air by first passing it through an active desiccant wheel to handle most of the latent load (humidity control), then post-cooling the resulting warm, dehumidified outdoor air as necessary to meet the temperature requirements of the conditioned space. However, this approach generally has not found market acceptance because of the relatively high purchase cost, high operational cost, large size and inefficiency of such systems.
When an active desiccant dehumidification wheel removes moisture from an air stream, heat is released as a result of the adsorption process in addition to the heat contained within the warm wheel media as it rotates from the hot regeneration air stream. The more moisture absorbed, the more heat released. This heat significantly increases the supply air temperature. In addition, removing large quantities of moisture from outdoor air (e.g., 60 grains) requires a high temperature air stream to regenerate the desiccant. In active desiccant wheels, this high regeneration temperature is supplied by an external heat source (e.g., a gas-fired heater). As mentioned, the heat imparted to the desiccant wheel further increases the supply air temperature. Based upon the literature for one of the best performing commercially available active desiccant wheels, a 60 grain reduction in outdoor air humidity would produce a 50xc2x0 F. increase in the outdoor air temperature. Herein lies a significant problem with the active desiccant preconditioning approach. If the desiccant wheel handles all or most of the outdoor air latent load, the amount of post cooling required to remove the sensible heat added by the dehumidification process will often be similar to that required to remove the humidity without the desiccant system. Consequently, this approach generally does not reduce the overall system energy consumption (total BTUs); rather, it increases it.
Another shortcoming of desiccant preconditioning approaches attempted heretofore is that such systems have required very large desiccant wheels to handle the significant latent load. For example, to process only 1500 cfm of outdoor air, active desiccant wheels as large as 42 inches have been applied. Including a standard cassette and drive assembly, the height and width of the wheel unit required is approximately 5 feet tall while a typical rooftop unit processing the same amount of air is only 33 inches tall. Most prior active desiccant systems have also employed a second, sensible only energy recovery wheel to mitigate much of the process heat gained as a result of the adsorption process. The size of the system required to accommodate these two wheels, regeneration and other components required is often four to five times the size of a comparable rooftop package unit. These large systems are particularly undesirable for commercial rooftop HVAC applications because they are more difficult to install, require greater structural reinforcement, and are less attractive. Architectural, engineering, economic and environmental considerations all drive the desire to reduce the size and weight of such packaged HVAC equipment.
Therefore, there is a significant need for energy-efficient, compact HVAC system that can effectively control the temperature and humidity of an indoor space while simultaneously providing high quantities of outdoor air to the space. The present invention provides these and other advantageous results.
The present invention provides systems and methods for controlling the temperature and humidity of air supplied to an enclosed space.
An apparatus of the present invention for dehumidifying air supplied by an air conditioning system includes a housing having a partition separating the interior of the housing into a supply portion and a regeneration portion. The supply portion has an inlet for receiving supply air from the leaving side of the air conditioning system cooling coil and an outlet for supplying air to the enclosed space. The regeneration portion has an inlet for receiving regeneration air and an outlet for discharging regeneration air. A fan in air flow communication with the regeneration portion creates a regeneration air stream.
The apparatus includes a rotatable desiccant wheel, which is preferably sized to handle approximately ⅓ of the air flow processed by the air conditioning system. The desiccant wheel preferably positioned substantially collinear or parallel to the partition such that a portion of the wheel extends into the supply portion and a portion of the wheel extends into the regeneration portion. The desiccant wheel rotates through the supply air stream and the regeneration air stream to dehumidify the supply air stream. The apparatus preferably includes a mechanism for varying the rotational speed of the desiccant wheel to control the amount of moisture removed from the supply air stream or heat transferred to the supply air stream. The apparatus also preferably includes a bypass damper between the inlet and the outlet side of the supply air portion around the active desiccant wheel for controlling the amount of supply air passing through the desiccant wheel. The bypass damper can also be modulated to accommodate varying outdoor air and desired supply air conditions by selectively bypassing the desiccant wheel.
A heat source (e.g., a direct-fired gas burner, indirect-fired burner or heating coil) warms the regeneration air stream as necessary to regenerate the desiccant wheel as it rotates through the regeneration air stream. Heated air that is a byproduct of an air conditioning system, a manufacturing process, and/or an electrical generation plant, for example, may also serve as the regeneration source. The apparatus can also include a duct or opening connecting the regeneration inlet air to the compartment that houses the air conditioning condenser to allow the regeneration heater inlet air to be preheated by the condenser coil.
The present invention also includes a hybrid air conditioning and dehumidifying apparatus and methods for using the apparatus to control the temperature and humidity of air supplied to an enclosed space. The hybrid unit includes a housing having a partition that separates the housing into a supply portion and a regeneration portion. The supply portion has an inlet for receiving air and an outlet for supplying air to the enclosed space. The regeneration portion has an inlet for receiving regeneration air and an outlet for discharging regeneration air. A fan in air flow communication with the regeneration portion creates the regeneration air stream and a fan in air flow communication with the supply portion creates the supply air stream. A cooling coil cools and/or dehumidifies the supply air stream. A bypass damper can be positioned in the supply section to allow a portion of the supply air leaving the cooling coil to bypass around the active desiccant wheel, preferably allowing approximately ⅓ of the supply air flow to pass through the desiccant wheel under normal operating conditions. A rotatable desiccant wheel positioned downstream of the cooling coil further dehumidifies the supply air stream. A portion of the desiccant wheel extends into the supply portion and a portion of the wheel extends into the regeneration portion, so that the wheel can rotate through the supply air stream and the regeneration air stream to exchange moisture between the air streams. A heat source heats the regeneration air stream as necessary to regenerate the desiccant wheel as it rotates through the regeneration air stream.